When a turbogenerator utilizes gaseous fuel to generate electricity, it is typically using natural gas from a natural gas pipeline. If the natural gas pipeline is in a residential or commercial area, the gas pressure is probably about two-tenths of a pound per square inch above atmospheric pressure (0.2 psig). The natural gas pipeline pressure is kept this low in residentially and commercially zoned areas for fire safety reasons. A line leak or line break at higher pressures could release massive amounts of natural gas into populated areas with the attendant risk of explosion and fire. In industrial locations, the natural gas pipeline pressure can be anywhere from twenty (20) psig to sixty (60) psig. Each natural gas pipeline has its own gas pressure standard. The utilities that supply these pipelines make little warranty of what that pressure will be or that it will be maintained at a relatively constant level.
A turbogenerator may need natural gas supplied to its combustor nozzle manifold at a pressure as low as one (1) psig when the turbogenerator is being started or at a pressure as high as forty (40) psig when the turbogenerator is being operated at full speed and at fill output power. One type of turbogenerator can operate and generate power at any speed between twenty-five thousand (25,000) rpm and one hundred thousand (100,000) rpm. Over this speed range, the gaseous fuel supply pressure requirements (in psig) can vary by twenty-five-to-one (25:1) and the gaseous fuel supply flow requirements can vary by twenty-to-one (20:1). The turbogenerator speed, combustion temperature and output power are controlled by the fuel pressure and the fuel flow rate established by the turbogenerator fuel control system. The pressure and flow of the gas delivered to the turbogenerator manifold must be precisely controlled (e.g. to within 0.01%) to adequately control the turbogenerator speed, combustion temperature and output power.
A fuel control system for a turbogenerator needs to be able to reduce the natural gas pressure when the turbogenerator is being started or operated at low speed and low output power. But this fuel control system must also be able to increase the natural gas pressure when the turbogenerator is operated at high speed and high output power. Thus, the fuel control system must have gas compression capability to increase the gas pressure. But it can reduce the gas pressure with either valves using Joule-Thompson expansion (which is wasteful of power) or with a turbine (which recovers the energy of the expanding natural gas and converts this into electrical power).
Most conventional gaseous fuel compression and control systems for turbogenerators utilize an oil lubricated reciprocating compressor driven by a three (3) phase, sixty (60) cycle induction motor to boost the natural gas pressure from whatever line pressure is available to a pressure of about one hundred (100) psig. There is typically an accumulator tank and a pressure sensor at the discharge of the reciprocating compressor. When the discharge pressure reaches about one hundred (100) psig, the pressure sensor turns the compressor motor off. The accumulator supplies the required turbogenerator gas flow when the compressor motor is turned off. The accumulator pressure decays with time until the pressure sensor determines that the pressure is below about sixty (60) psig, at which time it turns the compressor motor on again. Once again the pressure rises to about one hundred (100) psig at which point the pressure sensor turns the compressor motor off. This process of pressure ramp up and pressure decay (from sixty (60) psig to one hundred (100) psig and back) continues as long as the turbogenerator is in operation. The compressor/accumulator discharge gas pressure is too high and poorly regulated for direct use by a turbogenerator. This pressure is regulated downward to match the requirements of the turbogenerator by a very precise mass flow control valve.
The mass flow control valve typically has a mass flow sensor that is insensitive to gas pressure, gas density, or gas temperature. The valve is a servosystem in its own right, adjusting its internal electromechanical orifice to prevent accumulator pressure variations from affecting the mass flow rate of natural gas delivered to the turbogenerator and to assure that the mass flow delivered to the turbogenerator is that commanded by the turbogenerator computer. The turbogenerator computer monitors the output power currently being demanded of the turbogenerator by the electrical load, computes the required changes in turbogenerator speed and combustion temperature required to supply that power, limits turbogenerator speed and combustion temperature to safe levels, then computes the mass flow of fuel required to achieve the latest desired turbogenerator speed and combustion temperature. The mass flow control valve is then commanded to deliver this mass flow rate.
The use of oil lubricated reciprocating compressors, fixed speed compressor motors, accumulators, on-off pressure control, mass flow control valves, etc. in conventional fuel control systems results in numerous shortcomings:
The fuel control system can be as large and heavy as the turbogenerator it supplies with gas and controls. PA1 The fuel control system can be as expensive as the turbogenerator it supplies with gas and controls. PA1 An oil coalescing filter with an oil return to the compressor oil sump as well as a depth filter are needed at the discharge of the compressor to prevent oil from contaminating the natural gas lines leading to the turbogenerator as well as contaminating the turbogenerator's nozzles, combustor and catalyst. PA1 If the oil coalescing filter and depth filter allow oil vapors or oil droplets to reach the natural gas lines leading to the turbogenerator, oil condensation and coalescing will allow liquid oil to plug these lines resulting in severe turbogenerator speed surges. PA1 If the oil coalescing filter and depth filter allow oil vapors or oil droplets to reach the turbogenerator's nozzles or combustor, varnish build up will affect combustion adversely. PA1 If the oil coalescing filter and depth filter allow oil vapors or oil droplets to reach any catalyst used by the turbogenerator in its combustion or post combustion emissions control, the catalyst will be poisoned and will cease to function. PA1 The compressor needs periodic servicing to check its oil level, top off its oil level and to change its oil. The filters require checking and periodic replacement. PA1 Turning the compressor on and off to control accumulator tank pressure shortens the compressor and motor life. PA1 The rings, rotary seals and sliding surfaces of the compressor wear and thus limit compressor life. PA1 The rotary seals of some compressor types can leak natural gas, especially after the passage of time and accumulated wear. PA1 The compressor produces pressure pulsations each time its piston strokes. These pulsations have to be overcome by compressing the gas to a higher pressure than would otherwise be needed (wasting power) and by the use of an accumulator tank and a fast acting mass flow control valve having a very high gain servosystem. PA1 The compressor/accumulator discharge pressure ramps up and decays down as the compressor is turned on and turned off to control accumulator tank pressure. These pressure variations also have to be overcome by compressing the gas to a higher pressure than would otherwise be needed (wasting power) and by the use of an accumulator tank and a fast acting mass flow control valve having a very high gain servosystem. PA1 The accumulator tank is large, heavy, and of at least medium cost. PA1 The mass flow control valve is expensive, complicated, prone to calibration drift, prone to sealing problems, prone to particle contamination, prone to friction induced hysterisis problems for some versions, adversely affected by electrical noise emanating from the turbogenerator (due to its high servosystem gains), and is of questionable reliability. PA1 Compressing the natural gas to a pressure far above that needed by the turbogenerator in order to have enough pressure differential across the mass flow control valve for the valve to operate well is very wasteful of natural gas compression power. PA1 (a) simple, reliable design with only one rotating assembly; PA1 (b) stable, surge-free operation over a wide range of operating conditions (i.e. from full flow to no flow); PA1 (c) long life (e.g., 40,000 hours) limited mainly by their bearings; PA1 (d) freedom from wear product and oil contamination since there are no rubbing or lubricated surfaces utilized; PA1 (e) fewer stages required when compared to a centrifigal compressor; and PA1 (f) higher operating efficiencies when compared to a very low specific-speed (high head pressure, low impeller speed, low flow) centrifugal compressor.
To avoid the aforementioned shortcomings of most conventional gaseous fuel compression and control systems, it is necessary to use a rotary compressor that is not oil lubricated, does not have rubbing shaft seals, and does not have sliding surfaces. Centrifugal compressors do meet these requirements. However, centrifugal compressors operate best (with high efficiencies) when they have a high throughput flow rate and a low pressure rise relative to their tip speed. These operating conditions are characterized as high specific-speed conditions. Under these conditions, a centrifugal compressor can operate with an efficiency on the order of seventy-eight percent (78%). But the flow rate and pressure rise requirements for the compressor in the gaseous fuel compression and control system are for a low specific-speed compressor (low throughput flow rate and high pressure rise relative to the compressor's tip speed). A centrifugal compressor operating under these conditions would have an efficiency of less than twenty percent (20%). Under these conditions it would require a very large number of centrifugal compressors in series (e.g. ten (10)) to produce the same pressure rise for a given tip speed as could one (1) helical flow compressor. A helical flow compressor is an attractive candidate for this application. A helical flow turbine can perform the function of the mass flow control valve while additionally generating electrical power when the natural gas line pressure is greater than that needed by the turbogenerator. If a helical flow machine is used that can function as both a compressor and a turbine, the natural gas need only be compressed to forty (40) psig instead of to one hundred (100) psig, thereby saving power.
A helical flow compressor/turbine operating as a compressor is a high-speed rotating machine that accomplishes compression by imparting a velocity head to each fluid particle as it passes through the machine's impeller blades and then converting that velocity head into a pressure head in a stator channel that functions as a vaneless diffuser. While in this respect a helical flow compressor has some characteristics in common with a centrifugal compressor, the primary flow in a helical flow compressor is peripheral and asymmetrical, while in a centrifugal compressor, the primary flow is radial and symmetrical. The fluid particles passing through a helical flow compressor travel around the periphery of the helical flow compressor impeller within a generally horseshoe shaped stator channel. Within this channel, the fluid particles travel along helical streamlines, the centerline of the helix coinciding with the center of the curved stator channel. This flow pattern causes each fluid particle to pass through the impeller blades or buckets many times while the fluid particles are traveling through the helical flow compressor, each time acquiring kinetic energy. After each pass through the impeller blades, the fluid particles reenter the adjacent stator channel where they convert their kinetic energy into potential energy and a resulting peripheral pressure gradient in the stator channel. The multiple passes through the impeller blades (regenerative flow pattern) allows a helical flow compressor to produce discharge heads of up to fifteen (15) times those produced by a centrifugal compressor operating at equal tip speeds. A helical flow compressor operating at low specific-speed and at its best flow can have efficiencies of about fifty-five percent (55%) with curved blades and can have efficiencies of about thirty-eight percent (38%) with straight radial blades.
A helical flow compressor can be utilized as a turbine by supplying it with a high pressure working fluid, dropping fluid pressure through the machine, and extracting the resulting shaft horsepower with a generator. Hence the term "compressor/turbine" which is used throughout this application.
Among the advantages of a helical flow compressor or a helical flow turbine are:
The flow in a helical flow compressor can be visualized as two fluid streams which first merge and then divide as they pass through the compressor. One fluid stream travels within the impeller buckets and endlessly circles the compressor. The second fluid stream enters the compressor radially through the inlet port and then moves into the horseshoe shaped stator channel which is adjacent to the impeller buckets. Here the fluids in the two streams merge and mix. The stator channel and impeller bucket streams continue to exchange fluid while the stator channel fluid stream is drawn around the compressor by the impeller motion. When the stator channel fluid stream has traveled around most of the compressor periphery, its further circular travel is blocked by the stripper plate. The stator channel fluid stream then turns radially outward and exits from the compressor through the discharge port. The remaining impeller bucket fluid stream passes through the stripper plate within the buckets and merges with the fluid just entering the compressor/turbine.
The fluid in the impeller buckets of a helical flow compressor travels around the compressor at a peripheral velocity which is essentially equal to the impeller blade velocity. It thus experiences a strong centrifugal force which tends to drive it radially outward, out of the buckets. The fluid in the adjacent stator channel travels at an average peripheral velocity of between five (5) and ninety-nine (99) percent of the impeller blade velocity, depending upon the compressor discharge flow. It thus experiences an inertial force which is much less than that experienced by the fluid in the impeller buckets. Since these two inertial forces oppose each other and are unequal, the fluid occupying the impeller buckets and the stator channel is driven into a circulating or regenerative flow. The fluid in the impeller buckets is driven radially outward and "upward" into the stator channel. The fluid in the stator channel is displaced and forced radially inward and "downward" into the impeller bucket.
The fluid in the impeller buckets of a helical flow turbine travels around the turbine at a peripheral velocity which is essentially equal to the impeller blade velocity. It thus experiences a strong centrifugal force which would like to drive it radially outward if unopposed by other forces. The fluid in the adjacent stator channel travels at an average peripheral velocity of between one hundred and one percent (101%) and two hundred percent (200%) of the impeller blade velocity, depending upon the compressor discharge flow. It thus experiences a centrifugal force which is much greater than that experienced by the fluid in the impeller buckets. Since these two inertial forces oppose each other and are unequal, the fluid occupying the impeller buckets and the stator channel is driven into a circulating or regenerative flow. The fluid in the impeller buckets is driven radially inward and "upward" into the stator channel. The fluid in the stator channel is displaced and forced radially outward and "downward" into the impeller bucket.
While the fluid in either a helical flow compressor or helical flow turbine is traveling regeneratively, it is also traveling peripherally around the stator-impeller channel. Thus, each fluid particle passing through a helical flow compressor travels along a helical streamline, the centerline of the helix coinciding with the center of the generally horseshoe shaped stator-impeller channel.